This invention relates to a fluid displacement apparatus, and more particularly, to a fluid displacement apparatus of the scroll type.
Scroll type fluid displacement apparatus are well known in the prior art. For example, U.S. Pat. No. 801,182 discloses a device including two scroll members each having a circular end plate and a spiroidal or involute spiral element. These scroll members are maintained angularly and radially offset so that both spiral elements interfit to make a plurality of line contacts between their spiral curved surfaces, thereby sealing off and defining at least one pair of fluid pockets. Relative orbital motion of the two scroll members is effected by a rotating crank mechanism. This motion shifts the line contacts along the spiral curved surfaces and, therefore, the fluid pockets change in volume. The volume of the fluid pockets increases or decreases depending on the direction of the orbital motion. Therefore, the scroll type fluid apparatus is applicable to compress, expand, or pump fluids.
In these scroll type fluid displacement apparatus, compression, expansion or pumping of the fluid is achieved by the change of volume of the fluid pockets defined between the spiral elements. The fluid pockets are defined by the line contacts between interfitting spiral elements, and the axial contacts between the axial end surface of each spiral element and the inner end surface of the end plate of the opposing scroll member. As the orbiting scroll member orbits, the line contacts shift along the spiral curved surfaces of the spiral elements, and the axial contacts slide on the inner end surface of each end plate. Effective sealing of the fluid pockets in these moving areas of contact is essential for efficient operation of the apparatus.
Various techniques have been used in the prior art to resolve the sealing problem, in particular, that relating to axial sealing. In U.S. Pat. No. 3,994,636, incorporated herein by reference, a seal element is mounted in a groove in the axial end surface of each spiral element. An axial force urging means in each groove, such as spring, urges the seal element toward the facing end surface of the end plate, thereby effecting an axial seal.
Because the seal element disclosed in the above patent is urged toward the facing end surface of the end plate by a spring or other axial force urging mechanism, over period of time, wear occurs between the end surface of the seal element and the end plate of the scroll member, especially when a light weight alloy, such as an aluminum alloy, is used as a material for the scroll member.
One solution to these problems is disclosed in commonly assigned copending application Ser. No. 312,755, filed Oct. 9, 1981, and incorporated herein by reference. This application discloses an involute anti-wear plate disposed between the axial end surface of a spiral element and the inner end surface of the opposite end plate. The involute anti-wear plate covers the area of the surface of the end plate where the other spiral element makes axial contact during orbital motion. Excessive wear or abrasion of the end plate is thereby prevented.
In this arrangement, shown in FIGS. 1-3, the end plate 2 of one scroll member is formed with a hole 3 at its center portion for passage of the fluid. The hole 3 is generally formed by a simple and low-cost drilling or end milling operation, so that the hole is circular and is formed near the inner end portion of spiral element 6 adjacent its inner wall, as shown in FIG. 1. Involute plate 4 disposed on the end plate 2 must be formed with matching hole 5 which is aligned with the hole 3 of end plate 2. During relative orbital movement of the scroll members, the inner end portion of spiral element 6' sweeps over hole 3 (see FIG. 2). If seal element 7 extends nearly to the inner end of spiral element 6'--a design which affords optimum sealing--seal element 7 will be quickly worn by abrasion against the edge of hole 3. Sealing is therefore usually compromised for the sake of seal longevity by using a seal element which terminates well short of the inner end of spiral element 6'.
As described in the aforesaid copending application Ser. No. 312,755, relative orbital movement of the scroll members diminishes the size of intermediate fluid pockets 8, 8' (FIG. 2) until the line contacts near the inner ends of spiral elements 6, 6' are broken. At this point the central high pressure fluid space or pocket communicates with intermediate fluid pockets 8, 8', causing a back flow of high pressure fluid into pockets 8, 8'. This results in an increase of the re-expansion volume, and a consequent loss of volumetric efficiency. This phenomenon is inherent in the operation of a scroll type compressor, but its undesirable effects can be minimized by efficient design. Volumetric efficiency can be maximized by delaying as much as possible communication of the central high pressure fluid space with the intermediate fluid pockets 8, 8', i.e., by maximizing the crank angle at which this communication occurs. Communication of the central high pressure fluid space with fluid pockets 8, 8' also can occur when the inner end portion of spiral element 6' is completely over hole 3, allowing high pressure fluid to leak back into pocket 8 behind the outer surface 6a of spiral element 6'. This can occur before the line contacts near the inner ends of spiral elements 6, 6' are broken.
In a compressor, hole 3 is generally larger than it would be in other types of scroll apparatus. Spiral element 6' therefore encounters hole 3 earlier (i.e., at a smaller crank angle) than it would a smaller hole. Hence, leakage of high pressure fluid behind surface 6a of spiral element 6' occurs earlier than desired, resulting in a premature increase in the re-expansion volume and a loss of volumetric efficiency.
Referring to FIGS. 2 and 11, the compression cycle of fluid in one fluid pocket will be described. FIG. 11 shows the relationship in a scroll type compressor of fluid pressure in an intermediate fluid pocket (8) to drive shaft crank angle, and shows that one compression cycle is completed in this case at a crank angle of 4.pi.. The compression cycle begins with the outer end of each spiral element 6, 6' in contact with the opposite spiral element, the suction stroke having finished. The state of fluid pressure in the fluid pockets in shown at point A in FIG. 11. The volume of the fluid pockets is reduced and fluid is compressed by the revolution of the orbiting scroll member until the crank angle reaches 2.pi., which state is shown by point B in FIG. 11. In this ideal case, where the re-expansion volume is zero, the fluid pressure is consequently increased to the discharge pressure (which is a function of the resiliency of reed valve 9--FIG. 3) by revolution of the orbiting scroll member, as shown by curve B-C-E in FIG. 11.
Generally in a compressor, however, after passing point C in FIG. 11, the pressurized intermediate pair of fluid pockets 8, 8' adjacent the central high pressure space are simultaneously connected to one another and to the high pressure space, which is located at the center of both spiral elements. As shown in FIG. 2, the high pressure space communicates to a discharge chamber through valve 9. At this time, the fluid pressure in the connected fluid pockets 8, 8' rises slightly due to mixing of high pressure fluid with the fluid in the connecting fluid pockets. This state is shown at point D' in FIG. 11. The fluid in the high pressure space is further compressed by revolution of the orbiting scroll member until it reaches the discharge pressure. This state is shown at point E' in FIG. 11. When the fluid pressure in the high pressure space reaches the discharge pressure, the fluid is discharged to the discharge chamber. In this case, the pressure in the fluid pocket 8 rises at the midway point of the compression cycle, resulting in a compression power loss which is represented by the shaded area in FIG. 11 between curves CE and D'E'.